Variable valve system of internal combustion engine

ABSTRACT

A variable valve system of an internal combustion engine includes a first mechanism that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto and a second mechanism that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, wherein the first and second mechanisms force mechanically the intake and exhaust valves to take smaller and larger operation angles respectively when the conversion actuation forces are not applied to the first and second mechanisms respectively.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to variable valve systems of an internal combustion engine, and more particularly to the variable valve systems of a type which is able to improve the fuel consumption of the internal combustion engine by improving a fuel combustibility at engine starting in cold air.

2. Description of the Related Art

In order to clarify the present invention, one related art disclosed in Jidousha Gijutsu (Automotive Technology) Vol. 43, No. 8, pages 14 to 19, published in 1989 will be briefly described before commencing description of the present invention.

The variable valve system of the publication comprises an intake valve operation angle varying mechanism that varies an operating angle of intake valves of an internal combustion engine and an exhaust valve operation angle varying mechanism that varies an operating angle of exhaust valves of the engine.

Each of the intake and exhaust valve operation angle varying mechanisms functions to cause the intake (or exhaust) valves to mechanically take a smaller operation angle when a hydraulic pressure for driving the mechanism is not applied thereto due to, for example, engine stopping or the like.

However, since, in the variable valve system of the publication, both the intake and exhaust valves are forced to take the smaller operation angle mechanically at the engine stopping, both the intake and exhaust valves inevitably show their smaller operation angle at the time of starting (or cranking) of the engine.

Now, combustion cycle at the engine starting will be considered. When, after going over the exhaust top dead center, a piston goes into an intake stroke and thus the piston starts to move down, a corresponding exhaust valve closes the exhaust port pretty soon, which however causes difficulty in moving back the high temperature exhaust gas (viz., high temperature EGR gas) from the exhaust side to a combustion chamber of the cylinder. The difficulty tends to prevent the interior of the cylinder to be heated quickly and thus unstable combustion of air/fuel mixture in the combustion chamber is induced, which brings about the problem of exhaust emission.

SUMMARY OF THE INVENTION

Accordingly, the present invention is provided by taking the above-mentioned drawbacks of the known variable valve system of an internal combustion engine into consideration.

That is, according to the present invention, there is provided a variable valve system of an internal combustion engine, which comprises an intake valve operation angle varying mechanism that mechanically controls intake valves to take a smaller operation angle and an exhaust valve operation angle varying mechanism that mechanically controls exhaust valves to take a larger operation angle, so that combustion of air/fuel mixture in the combustion chamber is improved and thus fuel efficiency is increased.

In accordance with a first aspect of the present invention, there is provided a variable valve system of an internal combustion engine, which comprises an intake valve operation angle varying mechanism that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto; and an exhaust valve operation angle varying mechanism that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, wherein the intake valve operation angle varying mechanism forces mechanically the intake valve to take a smaller operation angle when receiving no conversion actuation force; and wherein the exhaust valve operation angle varying mechanism forces mechanically the exhaust valve to take a larger operation angle when receiving no conversion actuation force.

In accordance with a second aspect of the present invention, there is provided a variable valve system of an internal combustion engine, which comprises an intake valve operation angle varying mechanism that varies an operation angle of an intake valve by using a hydraulic force or electric force as a conversion actuation force; and an exhaust valve operation angle varying mechanism that varies an operation angle of an exhaust valve by using a hydraulic force or electric force as a conversion actuation force, wherein when receiving no conversion actuation force, the intake valve operation angle varying mechanism forces the intake valve to take and hold an operation angle smaller than that taken when receiving the conversion actuation force, and wherein when receiving no conversion actuation force, the exhaust valve operation angle varying mechanism forces the exhaust valve to take and hold an operation angle larger than that taken when receiving the conversion actuation force.

In accordance with a third aspect of the present invention, there is provided a variable valve system of an internal combustion engine, which comprises a first means that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto; and a second means that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, wherein the first means forces mechanically the intake valve to take a smaller operation angle when receiving no conversion actuation force and wherein the second means forces mechanically the exhaust valve to take a larger operation angle when receiving no conversion actuation force.

In accordance with a fourth aspect of the present invention, there is provided, in a variable valve system of an internal combustion engine which comprises an intake valve operation angle varying mechanism that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto and an exhaust valve operation angle varying mechanism that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, a method for controlling the variable valve system which comprise controlling the intake valve operation angle varying mechanism to force mechanically the intake valve to take a smaller operation angle when no conversion actuation force is applied to the intake valve operation angle varying mechanism; and, controlling the exhaust valve operation angle varying mechanism to force mechanically the exhaust valve to take a larger operation angle when no conversion actuation force is applied to the exhaust valve operation angle varying mechanism.

BRIEF DESCRIPTION OF THE DRAWINGS

Other objects and advantages of the present invention will become apparent from the following description when taken in conjunction with the accompanying drawings, in which:

FIG. 1 is a schematic view of a variable valve system of an internal combustion engine, which is a first embodiment of the present invention;

FIGS. 2A and 2B are sectional views of an intake side lift varying mechanism (viz., Intake VVL), in which FIG. 2A depicts an operation made when a smaller lift control is carried out, and FIG. 2B depicts an operation made when a larger lift control is carried out;

FIGS. 3A and 3B are sectional views of an exhaust side lift varying mechanism (viz., Exhaust VVL), in which FIG. 3A depicts an operation made when a smaller lift control is carried out, and FIG. 3B depicts an operation made when a larger lift control is carried out;

FIG. 4 is a valve lift characteristic diagram depicting the valve lift characteristic of the intake and exhaust side lift varying mechanisms of the first embodiment of the present invention;

FIG. 5 is a flowchart depicting controlled operation steps carried out by a controller used in the first embodiment of the present invention;

FIGS. 6A and 6B are sectional views of a phase varying mechanism used in an exhaust side (and intake side) in a second embodiment of the present invention, in which the views are taken along the line A-A of FIG. 7;

FIG. 7 is a longitudinal sectional view of the phase varying mechanism of the second embodiment;

FIG. 8 is a valve lift characteristic diagram depicting the valve lift characteristic of the intake and exhaust side phase varying mechanisms of the second embodiment of the present invention;

FIG. 9 is a flowchart depicting controlled operation steps carried out by a controller used in the second embodiment of the present invention;

FIG. 10 is a view similar to FIG. 1 but showing a variable valve system of an internal combustion engine, which a third embodiment of the present invention;

FIGS. 11A and 11B are sectional views of an intake side lift varying mechanism of the third embodiment, in which FIG. 11A depicts a minimum lift condition, and FIG. 11B depicts a maximum lift condition;

FIGS. 12A and 12B are sectional views of an electric actuator used in the third embodiment, in which FIG. 12A depicts an intake valve operation mode made when the minimum lift control is carried out, and FIG. 12B depicts an intake valve operation mode made when the maximum lift control is carried out;

FIG. 13 is a valve lift characteristic diagram depicting the valve lift characteristic of the intake and exhaust side lift varying mechanisms of the third embodiment of the present invention; and

FIG. 14 is a flowchart depicting controlled operation steps carried out by a controller used in the third embodiment of the present invention.

DETAILED DESCRIPTION OF THE INVENTION

In the following, embodiments of a variable valve system of the present invention will be described in detail with reference to the accompanying drawings.

In the illustrated embodiments, a four-cycle multi cylinder gasoline internal combustion engine is used.

First Embodiment

In FIG. 1, there is shown a variable valve system of a first embodiment of the present invention. The internal combustion engine to which the variable valve system is practically applied has two intake valves 1 and 1 and two exhaust valves 2 and 2 for each cylinder.

As is seen from FIG. 1, for controlling the two intake valves 1 and 1, there is employed an intake valve operation angle varying mechanism (or intake VVL) 3, and for controlling the two exhaust valves 2 and 2, there is employed an exhaust valve operation angle varying mechanism (or exhaust VVL) 4. Each mechanism 3 or 4 is constructed to stepwisely vary a valve operation angle and a valve lift amount in accordance with an operation condition of the engine.

First, the detail of the intake VVL 3 will be described.

That is, as is seen from FIGS. 1 and 2A, the intake VVL 3 comprises an intake camshaft 5 that extends in a longitudinal direction of the engine, a larger lift cam 7 integrally mounted on the intake camshaft 5 at a position through which an axis of the corresponding cylinder passes, two smaller lift cams 8 and 8 integrally mounted on the intake camshaft 5 at both sides of the larger lift cam 7, a rocker shaft 9 located obliquely below the intake camshaft 5, a main rocker arm 10 including two side followers 10 a and 10 a that are pivotally supported by the rocker shaft 9 at positions just below the smaller lift cams 8 and 8 and have respective leading ends contactable with stem ends of the two intake valves 1 and 1, a sub-rocker arm 11 having a center follower 11 a at a position below the larger lift cam 7 and being subjected to a so-called “lost motion”, a lost-motion mechanism 12 (see FIG. 2A) installed in the sub-rocker arm 11 to bias the sub-rocker arm 11 toward the larger lift cam 7, a supporting shaft 13 (see FIG. 2A) fixed to the main rocker arm 10, a lever member 14 (see FIG. 2A) that is pivotally supported by the supporting shaft 13 and constructed to induce a synchronous swing of the sub-rocker arm 11 and the main rocker arm 10 when engaged with a lower stepped part 11 b of the sub-rocker arm 11 and cancel the synchronous swing of the arms 11 and 10 when disengaged from the lower stepped part 11 b, a hydraulic plunger 15 for producing a force in a direction to bring about the engagement between the lever member 14 and the lower end 11 b of the sub-rocker arm 11, and a return spring 16 for producing a force in a direction to cancel the engagement between the lever member 14 and the lower end 11 b of the sub-rocker arm 11. It is thus to be noted that when the engagement between the lever member 14 and the lower stepped part 11 b of the sub-rocker arm 11 is kept cancelled, so-called “lost motion” is applied to the sub-rocker arm 11 relative to the main rocker arm 10.

As is seen from FIG. 2A, the sub-rocker arm 11 is pivotally supported by a supporting shaft 23 provided at a rear end portion of the main rocker arm 10.

As is seen from FIG. 2A, the lost-motion mechanism 12 comprises a plunger 25 that is slidably received in a bore formed in a lower part of the sub-rocker arm 11, a projection 10 b that is formed on a central lower portion of the main rocker arm 10 and contactable with the plunger 25 and a lost motion spring 24 that is compressed between the plunger 25 and a bottom of the bore of the sub-rocker arm 11 to bias the plunger 25 toward the projection 10 b. In case where the lever member 14 is not engaged with the lower stepped part 11 b of the sub-rocker arm 11, the sub-rocker arm 11 is forced to make a lost motion thereof relative to the main rocker arm 10 through the plunger 25 and the lost motion spring 24.

As will be understood from FIG. 2A, the hydraulic plunger 15 is moved forward and back by a hydraulic pressure that is charged into (or discharged from) a hydraulic chamber 17 from an oil pump 19 (see FIG. 1) through a first electromagnetic switch valve 20.

That is, into or from the hydraulic chamber 17, there is charged or discharged a hydraulic pressure through hydraulic pressure passages 18 a and 18 b respectively formed in the rocker shaft 9 and the main rocker arm 10, as is seen from FIG. 2A. As is seen from FIG. 1, the first electromagnetic switch valve 20 is arranged to switch a connection between one of a passage led to the hydraulic pressure passages 18 a and 18 b, a drain passage 21 and a discharge passage 19 a of the oil pump 19.

Accordingly, when the first electromagnetic switch valve 20 is applied with ON signal (viz., energizing signal) and thus hydraulic pressure is fed to the hydraulic chamber 17 from the oil pump 19, the hydraulic plunger 15 is projected causing a leading end of the lever member 14 to engage with the lower stepped part 11 b of the sub-rocker arm 11. Upon this, the intake valves 1 and 1 are controlled to show a larger lift mode, as is indicated by LI2 in FIG. 4, following the a cam profile surface of the larger lift cam 7 by the swing movement of the sub-rocker arm 11 through the main rocker arm 10.

While, when the first electromagnetic switch valve 20 is applied with OFF signal (viz., de-energizing signal) and thus the hydraulic pressure is not fed to the hydraulic chamber 17, the hydraulic plunger 15 is moved back by the force of the return spring 16 releasing the leading end of the lever member 14 from the lower stepped part 11 b of the sub-rocker arm 11 cancelling the engagement therebetween. Upon this, the sub-rocker arm 11 is permitted to make the lost motion due to the work of the lost motion mechanism 12, and thus, the intake valves 1 and 1 are controlled to show a smaller lift mode, as is indicated by LI1 in FIG. 4, following cam profile surfaces of the identical smaller lift cams 8 and 8 by the swing movement of the main rocker arm 10.

Now, a case wherein, due to engine stop or the like, the hydraulic pressure (viz., mode changing force) of the oil pump 19 is not produced will be considered. In this case, the intake valves 1 and 1 are mechanically controlled to take the smaller lift mode irrespective of ON/OFF positions of the first electromagnetic switch valve 20 (which is called “default position”). That is, in the case, the two intake valves 1 and 1 are forced to take a default condition in the smaller lift mode.

[Exhaust VVL]

As is mentioned hereinabove, the exhaust VVL 4 is substantially the same in construction as the intake VVL 3. However, in the exhaust VVL 4, an application direction of the hydraulic pressure against the hydraulic plunger is different from that in the intake VVL 3.

Now, the exhaust VVL 4 will be described in the following with reference to FIGS. 1 and 3A. For ease of understanding, reference numerals indicating parts and portions shown in FIG. 3 are made different from those corresponding to the parts and potions shown in FIG. 2.

As is seen from FIGS. 1 and 3A, the exhaust VVL 4 comprises an exhaust camshaft 6 that extends in parallel with the above-mentioned intake camshaft 5, a larger lift cam 31 integrally mounted on the exhaust camshaft 6 at a position through which an axis of the corresponding cylinder passes, two smaller lift cams 32 and 32 integrally mounted on the exhaust camshaft 6 at both sides of the larger lift cam 31, a rocker shaft 33 located obliquely below the exhaust camshaft 6, a main rocker arm 34 including two side followers 34 a and 34 a that are pivotally supported by the rocker shaft 33 at positions just below the smaller lift cams 32 and 32 and have respective leading ends contactable with stem ends of the two exhaust valves 2 and 2, a sub-rocker arm 35 having a follower 35 a at a position below the larger lift cam 31 and being subjected to a “lost motion”, a lost-motion mechanism 36 (see FIG. 3A) installed in the sub-rocker arm 35 to bias the sub-rocker arm 35 toward the larger lift cam 31, a supporting shaft 41 (see FIG. 3A) fixed to the main rocker arm 34, a lever member 38 (see FIG. 3A) that is pivotally supported by the supporting shaft 41 and constructed to induce a synchronous swing of the sub-rocker arm 35 and the main rocker arm 34 when engaged with a lower stepped part 35 a of the sub-rocker arm 35 and cancel the synchronous swing of the arms 35 and 34 when disengaged from the lower stepped part 35 a, a hydraulic plunger 39 for inducing the engagement/disengagement of the lever member 38 and a return spring 40 installed in the main rocker arm 34. It is thus to be noted that when the engagement between the lever member 38 and the lower stepped part 35 a of the sub-rocker arm 35 is kept cancelled, a so-called “lost motion” is applied to the sub-rocker arm 35 relative to the main rocker arm 34.

As is seen from FIG. 3A, the sub-rocker arm 35 is pivotally supported by a supporting shaft 41 provided at a rear end portion of the main rocker arm 34.

As is seen from FIG. 3A, the lost-motion mechanism 36 comprises a plunger 42 that is slidably received in a bore formed in a lower part of the sub-rocker arm 35, a projection 34 b that is formed on a central lower portion of the main rocker arm 34 and contactable with the plunger 42 and a lost motion spring 43 (see FIG. 3B) that is compressed between the plunger 42 and a bottom of the bore of the sub-rocker arm 35 to bias the plunger 42 toward the projection 34 b. In case where the lever member 38 is not engaged with the lower stepped part 35 a of the sub-rocker arm 35, the sub-rocker arm 35 is forced to make a lost motion thereof relative to the main rocker arm 34 through the plunger 42 and the lost motion spring 43.

As will be understood from FIG. 3A, the hydraulic plunger 39 is moved backward (leftward in the drawing) in a bore formed in the main rocker arm 34 when a hydraulic pressure (viz., mode changing force) is fed from the oil pump 19 (see FIG. 1) to a hydraulic chamber 44 through a second electromagnetic switch valve 47. As shown, the hydraulic chamber 44 is defined between the bore formed in a central lower portion of the main rocker arm 34 and the hydraulic plunger 39 that is slidably received in the bore. While, when, due to OFF operation of the second electromagnetic switch valve 47, the hydraulic pressure is not fed to the hydraulic chamber 44, the hydraulic plunger 39 is forced to project toward the lever member 38 by a force of a coil spring 46 compressed between the hydraulic plunger 39 and a fixed part of the bore of the main rocker arm 34.

That is, into or from the hydraulic chamber 44, there is introduced or discharged the hydraulic pressure through hydraulic pressure passages 45 a and 45 b formed in the rocker shaft 33 and the main rocker arm 34. As is seen from FIG. 1, the second electromagnetic switch valve 47 is arranged to switch a connection between one of a passage led to the hydraulic pressure passages 45 a and 45 b, a drain passage 48 and a discharge passage 19 a of the oil pump 19.

Accordingly, when, due to energization of the second electromagnetic switch valve 47, the hydraulic pressure is fed to the hydraulic chamber 17 from the oil pump 19, the hydraulic plunger 39 is forced to move backward (viz., leftward in FIG. 3A) against the force of the coil spring 46 causing the leading end of the lever member 38 to release from the lower stepped part 35 a of the sub-rocker arm 35. Upon this, due to the work of the lost motion mechanism 36, the sub-rocker arm 35 is brought into a lost motion condition, and thus, the exhaust valves 2 and 2 are controlled to show a smaller lift mode, as is indicated by LE1 in FIG. 4, following cam profile surfaces of the identical smaller lift cams 32 and 32 by the swing movement of the main rocker arm 34.

While, when the second electromagnetic switch valve 47 is applied with OFF signal (viz., de-energizing signal) and thus the hydraulic pressure is not fed to the hydraulic chamber 44 from the oil pump 19, the hydraulic plunger 39 is forced to project forward (viz., rightward in FIG. 3B) due to the force of the coil spring 46 thereby pressing a lower end of the lever member 38. Upon this, the leading end of the lever member 38 is brought into engagement with the lower stepped part 35 a of the sub-rocker arm 35 as is seen from FIG. 3B. Thus, the exhaust valves 2 and 2 are controlled to show a larger lift mode, as is indicated by LE2 in FIG. 4, following the cam profile surface of the larger lift cam 31 by the swing movement of the sub-rocker arm 35 through the main rocker arm 34.

When, due to engine stop or the like, the hydraulic pressure is not produced by the oil pump 19, the exhaust valves 2 and 2 are mechanically controlled to take the larger lift mode irrespective of ON/OFF positions of the second electromagnetic switch valve 47 (which is called “default position”). That is, in this case, the two exhaust valves 2 and 2 are forced to take a default condition in the larger lift mode.

It is to be noted that the default condition of the exhaust valves 2 and 2 appears in the larger lift mode induced when, due to engine stop of the like, the hydraulic pressure is not produced by the oil pump 19 and the afore-mentioned default condition of the intake valves 1 and 1 appears in the smaller lift mode induced when, due to engine stop or the like, the hydraulic pressure is not produced by the oil pump 19. In other words, in the Exhaust VVL4, the default condition appears in the larger lift mode, while in the Intake VVL3, the default condition appears in the smaller lift mode.

The controller (ECU) 22 receives information signals from various sensors, such as crank-angle sensor, airflow meter, engine cooling water temperature sensor, oil temperature sensor, throttle open degree sensor, etc., for calculating or detecting a current engine operation condition and controlling the first and second electromagnetic switch valves 20 and 47 based on the calculated current engine operation condition.

Effects of the First Embodiment

As has been mentioned hereinabove, in the default condition, the intake valves 1 and 1 show a smaller lift (LI1) and a smaller operation angle and the exhaust valves 2 and 2 show a larger lift (LE2) and a larger operation angle, as is seen from FIG. 4. That is, when, under engine stop and/or during cranking of the engine, a sufficient hydraulic pressure is not produced, the intake and exhaust valves 1, 1, 2 and 2 take a mechanically stable default condition.

It is to be noted that a default valve timing thus induced from such mechanically stable default condition of the valves 1, 1, 2 and 2 provides the engine with a stable combustion at its starting, particularly at the cold starting, which can reduce emission, such as HC and the like, from the exhaust gas of the engine.

It is further to be noted that since the engine already takes such stable default valve timing under the engine stop, the emission reduction is effected at the initial stage of combustion of the engine.

The exhaust valves 2 and 2 are controlled to show the larger operation angle DE2 (lift=LE2). Thus, even when the piston comes into an intake stroke after passing through the exhaust top dead center and thus the piston starts to move downward, closing of the exhaust ports by the exhaust valves 2 and 2 is not immediately made and thus, a larger amount of high temperature combustion gas (viz., high temperature EGR gas) can be reversely led into the corresponding cylinder from the exhaust ports for a relatively long time (viz., a first period indicated by (1)-2 in FIG. 4) from the time of the exhaust top dead center (TDC) to the time of the exhaust valve close timing (EVC2).

While, the intake valves 1 and 1 are controlled to show the smaller operation angle (DI1), and thus, the period (viz., a second period indicated by (2)-1 in FIG. 4) from the time of intake valve opening timing (IVO1) to the time of the exhaust top dead center (TDC) is relatively short. Accordingly, the amount of the high temperature combustion gas (viz., high temperature EGR gas) that is produced at a later stage of the exhaust stroke, returned back toward the lower temperature intake ports and intake pipe through the intake valves 1 and 1 during the second period and thus cooled to become so-called low temperature EGR gas can be reduced, and thus, the amount of the low temperature EGR gas led again into the corresponding cylinder in a subsequent intake stroke can be reduced.

As will be understood from the above, in such case, in the amount of EGR gas led into the cylinder, the ratio of the high temperature EGR gas provided by the first period can be increased while reducing the ratio of the low temperature EGR gas provided by the second period.

Furthermore, since the exhaust valves 2 and 2 are controlled to show a larger operation angle, these valves show relatively early open timing (EVO2), and thus the combustion gas is discharged to the exhaust system while the temperature of the gas is not sufficiently lowered. This means that the catalyst placed at a downstream part of the exhaust system can be effectively heated promoting activity of the catalyst. That is, exhaust emission can be reduced.

Furthermore, since the intake valves 1 and 2 are controlled to show a smaller operation angle DI1 (lift=LI1), the closing timing of the intake valves 1 and 1 is advanced near the intake bottom dead center (BDC), and thus, effective compression ratio can be increased, which promotes the combustion stability.

Due to the effect of increasing the ratio of the high temperature EGR gas and the effect of increasing the effective compression ratio, the starting combustion of the engine can be stabilized and the activity of the catalyst can be increased. Thus, the exhaust emission, such as HC, etc., from the engine can be effectively reduced.

Furthermore, since the intake valves 1 and 1 and exhaust valves show the smaller and larger operation angles respectively just at the initial period of engine starting, the above-mentioned advantageous effects are assuredly obtained just at the initial period of engine starting.

Referring to FIG. 5, there is shown a flowchart that depicts controlled operation steps executed by the controller 22 for controlling the variable valve system of the first embodiment of the present invention.

At step S1, judgment is carried out as to whether a current state is an engine stopping condition or not, that is, whether an engine key has returned to a key-off position or not.

If YES, the operation flow goes to step S2 where an engine stop signal (or fuel cut signal) is issued. Upon this, the engine speed Ne is gradually lowered. Due to lowering of the engine speed, the hydraulic pressure from the oil pump 19 is gradually lowered, and finally, the hydraulic pressure fails to have a mode changing force irrespective of the control positions of the first and second electromagnetic switching valves 20, 47.

Thus, at step S3, due to the reasons as mentioned hereinabove, the intake and exhaust valves 1, 1, 2 and 2 are moved to their default positions. That is, the intake valves 1 and 1 take the default but stable position ready for the control with the smaller operation angle, and the exhaust valves 2 and 2 take the default but stable position ready for the control with the larger operation angle.

At step S4, operation of the engine is stopped.

If NO at step S1, that is, if the current state is not the engine stop condition, the operation flow goes to RETURN.

At step S5, judgment is carried out as to whether the current state is an engine start condition or not, that is, whether the engine key is turned to a key-on position or not.

If YES, the operation flow goes to step S6 where, for assuring the positioning of the intake and exhaust valves 1, 1, 2 and 2 at the default positions, control signals are fed to the first and second electromagnetic switch valves 20, 47 for forcing the intake and exhaust valves 1, 1, 2 and 2 to take the above-mentioned default positions. The control of this step S6 is aimed to make an effort to move the intake and exhaust valves 1, 1, 2 and 2 to the default positions with the reduced hydraulic pressure from the pump 19. Actually, subsequent cranking of the engine tends to induce a fluctuation of the hydraulic pressure from the pump 19, and thus it is desired to move the valves 1, 1, 2 and 2 to the default positions even though the hydraulic pressure from the pump 19 is small. But, if not necessary, the control of step S6 may be cancelled.

If NO at step S5, that is, when the engine key is not turned to the key-on position, the operation flow goes to RETURN.

Then at step S7, a cranking signal is issued for cranking the engine.

Then, at step S8, judgment is carried out as to whether the engine cranking speed has reached to a predetermined speed or not. If YES, that is, when the cranking speed has reached to the predetermined speed that assures a perfect ignition in the engine, the operation flow goes to step S9.

At step S9, fuel injection and ignition are applied to the engine for starting the engine. This will be called “perfect ignition control” hereinafter.

If NO at step S8, the operation flow goes back to the entrance of the same step S8.

It is to be noted that at the time when step S9 is carried out, the intake and exhaust valves 1, 1, 2 and 2 are in the above-mentioned default but stable positions, and thus, the combustion for starting the engine can be stable and emission at the engine starting can be reduced.

At step S10, judgment is carried out as to whether a predetermined time has passed from the engine starting or not. If YES, the operation flow goes to step S11. If NO, the operation flow goes back to the entrance of the same step S10.

At step S11, judgment is carried out as to whether the engine temperature T is lower than a predetermined temperature TO or not. If YES, the operation flow goes to RETURN. However, if NO, the operation flow goes to step S12.

At step S12, instruction signals are fed to the exhaust VVL 4 to reduce the operation angle of the exhaust valves 2 and 2 to the smaller operation angle LD1 (smaller lift LE1). This is because if the engine temperature exceeds the predetermined temperature TO, it is so judged that the warming-up operation of the engine has sufficiently progressed. That is, under such condition, abnormal combustion causing knocking, pre-ignition and the like tends to appear because of the higher temperature of the engine. In the default valve timing, a large amount of high temperature EGR gas is led into the cylinder, and thus the cold combustion of the engine is improved. However, when the engine temperature is increased, the abnormal combustion tends to appear by the presence of the high temperature EGR.

When now the operation angle (or lift) of the exhaust valves 2 and 2 is reduced, the first period (viz., the period indicated by (1)-1 in FIG. 4) from the time of the exhaust top dead center (TDC) to the time of the exhaust valve close timing (EVC1) is reduced, and thus, the amount of the high temperature gas (viz., high temperature EGR gas) that is reversely led into the corresponding cylinder from the exhaust port side through the exhaust valves 2 and 2 is reduced. With this reduction, excessive heating in the cylinder is suppressed, and thus, the abnormal combustion causing knocking, pre-ignition and the like can be avoided.

Referring back to the flowchart of FIG. 5, at step S13, judgment is carried out as to whether the engine speed Ne has reached to a predetermined speed NO or not. If YES, the operation flow goes to step S14. If NO, the same judgment is carried out after waiting for a given time.

At step S14, instruction signals are fed to the intake VVL 3 to increase the operation angle of the intake valves 1 and 1 to the larger operation angle DI2 (larger lift LI2). This is because if the operation angle of the intake valves 1 and 1 is kept at the smaller operation angle, the lift is kept low and the closing timing IVC1 of the intake valves 1 and 1 is kept near the bottom dead center, and thus, a sufficient charging efficiency at the high engine speed range is not obtained. In such case, sufficient torque and output are not produced by the engine. Accordingly, at the step S14, the operation angle of the intake valves 1 and 1 is increased to the larger operation angle DI2 raising the lift to LI2, so that the valve close timing is retarded to the time of IVC2 to increase the charging efficiency, torque and output.

While, when the torque and output are increased, the thermal load is also increased, which increases a possibility of abnormal combustion such as knocking and pre-ignition. However, in the larger operation angle of the intake valves 1 and 1, the valve open timing IVO is advanced. In other words, the second period from the time of the valve open timing IVO2 to that of the exhaust top dead center (TDC) is expanded, and thus, the rate of the lower temperature EGR returned back to the cylinder is relatively increased thereby restraining the abnormal combustion. Accordingly, by the action of step S14, the undesired knocking and pre-ignition can be restrained while assuring the torque and output in the high engine speed range.

It is to be noted that the above-mentioned stable combustion and emission reduction at the engine starting induced by the default valve timing have a mechanical fail safe function by which such stable combustion and emission reduction are still assured even if the electromagnetic switch valves 20 and 47 go down due to break in the cable or the like. That is, if, due to the cable break or the like, the electromagnetic switch valves 20 and 47 become constantly fed with OFF signal (viz., de-energized) and thus the valves 20 and 47 are forced to take OFF positions, or if, due to malfunction of an electronic control system, ON signal is erroneously fed to the valves 20 and 47 causing them to take their ON positions, the condition or mode of the default valve timing is stable (viz., not change) at the engine starting when the hydraulic pressure from the oil pump 19 is not sufficiently fed to the Intake/Exhaust VVLs 3 and 4.

Second Embodiment

In FIGS. 6A, 6B and 7, there is shown a variable valve system of a second embodiment of the present invention.

As will become apparent as the description proceeds, the variable valve system of this second embodiment generally comprises means corresponding to the Intake VVL3 of the first embodiment, means corresponding to the Exhaust VVL4 of the first embodiment, an Intake VTC 49 that is an intake phase varying mechanism and an Exhaust VTC 50 that is an exhaust phase varying mechanism.

It is to be noted that the Intake VTC 49 and the Exhaust VTC 50 are substantially the same in construction except that respective camshafts have different phase conversion angles θI and θE.

Thus, for simplification of description, the following explanation will be directed to only the Exhaust VTC 50 and made with reference to FIGS. 6A, 6B and 7.

As is seen from FIGS. 6A and 7, the Exhaust VTC 50 is of a vane type that generally comprises a timing sprocket 51 that is driven by a crankshaft of the engine for driving the exhaust camshaft 6, a vane member 52 that is connected to an end of the exhaust camshaft 6 and rotatably received in the timing sprocket 51, and a hydraulic circuit 53 that rotates the vane member 52 in normal and reverse directions with the aid of the hydraulic pressure.

The timing sprocket 51 comprises a housing 54 that has the vane member 52 rotatably received therein, a circular front cover 55 that covers a front opening of the housing 54 and a generally circular rear cover 56 that covers a rear opening of the housing 54. The housing 54, the front cover 55 and the rear cover 56 are united together by means of four connecting bolts 79. As is seen from FIG. 6A, each connecting bolt 79 extends along an axis of the exhaust camshaft 6.

As is seen from FIGS. 6A and 7, the housing 54 is a cylindrical member whose front and rear ends are opened. Four shoes (or partition walls) 54 a are formed on a cylindrical inner surface of the housing 54 at generally 90° intervals. As shown, each shoe 54 a projects radially inward.

As is seen from FIG. 6A, each shoe 54 a has a generally trapezoidal shape and has at a generally center portion an opening 54 b through which corresponding one of the connecting bolts 79 passes. As is seen from the drawing, an inner end of each shoe 54 a is formed with a cut for slidably receiving therein a seal member 58. Although not shown in the drawing, a leaf spring is arranged to bias the seal member 58 radially inward, that is, toward the vane rotor 52 a.

The front cover 55 is shaped circular and has, as is seen FIG. 7, a larger support opening 55 a at its center portion. Although not shown in the drawings, the front cover 55 is formed at its peripheral portion with four bolt openings through which the connecting bolts 79 pass.

As is seen from FIG. 7, the rear cover 56 is formed at its rear end with a gear portion 56 a to which a timing chain is meshed and the rear cover 56 is formed at its generally center part with a larger opening 56 b.

As is seen from FIGS. 6A and 7, the above-mentioned vane member 52 comprises an annular vane rotor 52 a that has at its center part a bolt opening, and four vanes 52 b that are integrally formed on a peripheral portion of the vane rotor 52 a at generally 90° intervals.

As is seen from FIG. 7, the vane rotor 52 a has both a smaller diameter front end rotatably received in the support opening 55 a of the front cover 55 and a smaller diameter rear portion rotatably received in the opening 56 b of the rear cover 56.

As will be understood from FIGS. 6A and 7, the vane member 52 is fixed to a front end of the exhaust camshaft 6 through a connecting bolt 57 that passes through aligned bores respectively formed in the vane rotor 52 a and the exhaust camshaft 6.

As is seen from FIG. 6A, three of the vanes 52 b are shaped generally rectangular and one of them is shaped generally trapezoidal. The three vanes 52 b are generally the same in size and shape, while the remaining one is larger than the vanes 52 b, as shown. With these four vanes 52 b, the vane member 52 has a satisfied weight balance.

As is seen from FIG. 6A, each vane 52 b is arranged between two neighboring shoes 54 a and has at its outer end a holding cut for slidably receiving therein a seal member 60. Although not shown in the drawing, a leaf spring is arranged to bias the seal member 60 radially outward, that is, toward a peripheral part of the housing 54, as shown.

As is seen from FIG. 6A, each of the vanes 52 b is formed, at one side thereof facing against the rotation direction of the exhaust camshaft 6, with two circular recesses 52 c.

As will be understood from FIG. 6B, by the four vanes 52 b and the four shoes 54 a which are alternately mated, there are defined four advancing hydraulic pressure chambers 61 and four retarding hydraulic pressure chambers 62.

As is seen from FIG. 7, the above-mentioned hydraulic circuit 53 includes a first hydraulic pressure passage 63 through which a hydraulic pressure is led into or discharged from the four advancing hydraulic pressure chamber 61, and a second hydraulic pressure passage 64 through which a hydraulic pressure is led into or discharged from the four retarding hydraulic pressure chamber 62.

These first and second hydraulic pressure passages 63 and 64 are connected to the discharge passage 19 a of the oil pump 19 and a drain passage 66 through a third electromagnetic switch valve 67. Designated by numeral 65 is a passage leading from the discharge passage 19 a to the switch valve 67. As is seen from the drawing, the oil pump 19 functions to pump oil from an oil pan 01 toward the passage 65, and the drain passage 66 has one end immersed in the oil in the oil pan 01.

As is seen from FIG. 7, the first and second hydraulic pressure passages 63 and 64 are formed in a cylindrical rod member 59. The cylindrical rod member 59 has a rear end portion liquid-tightly received in a cylindrical bore 52 d formed in the vane rotor 52 a, as shown. A front end portion of the cylindrical rod member 59 has the switch valve 67 connected thereto.

Between an outer surface of the cylindrical rod member 59 and an inner surface of the cylindrical bore 52 d, there are disposed three annular sealing members 70 for sealing the first and second hydraulic pressure passages 63 and 64. More specifically, due to provision of the sealing members 70, the two pressure passages 63 and 64 are hermetically isolated from each other.

The first hydraulic pressure passage 63 includes an oil chamber 63 a that is defined in the cylindrical bore 52 d at a deep part facing the rear end of the cylindrical rod member 59 and four branch passages 63 b that are formed in the vane rotor 52 a and extend radially outward from the oil chamber 63 a to the four advancing hydraulic pressure chambers 61 respectively.

While the second hydraulic pressure passage 64 includes an annular groove 64 a that is formed on an outer surface of the rear end portion of the cylindrical rod member 59 and a second hydraulic passage 64 b that is shaped L and formed in the vane rotor 52 a to connect the annular groove 64 a to the four retarding hydraulic pressure chambers 62.

As is seen from FIG. 7, the third electromagnetic switch valve 67 is of a four-port-three-position type that has a valve body. A coil spring (no numeral) is installed in the valve 67 to bias the valve body in a given direction. Upon operation, the valve body is moved to selected one of three positions which are a first position “I” as shown where the first hydraulic pressure passage 63 is connected to the drain passage 66 and the second hydraulic pressure passage 64 is connected to the passage 65 from the pump 19, a second position “II” where the above-mentioned connections are both cancelled and a third position “III” where the above-mentioned connections are reversed as shown. These first, second and third positions “I”, “II” and “III” are switched by the controller 22.

When a control current is not applied to the third switch valve 67, the switch valve 67 takes the first position “I” as shown in FIG. 7. In this case, the passage 65 is connected to the retarding hydraulic pressure chambers 62 through the second pressure passage 64 and the drain passage 66 is connected to the advancing hydraulic pressure chambers 61 through the first pressure passage 63.

In case of Intake VTC 49, a fourth electromagnetic switch valve 68 is used. However, since the fourth switch valve 68 is substantially the same as the third switch valve 67 for Exhaust VTC 50, detailed explanation of the fourth switch valve 68 will be omitted.

The controller 22 is substantially the same as the controller 22 used in the above-mentioned first embodiment. As is seen from FIG. 1, in the first embodiment, Intake VVL 3 (first electromagnetic switch valve 20) and Exhaust VVL 4 (second electromagnetic switch valve 47) are controlled by the controller 22.

Referring back to FIG. 7, in the second embodiment, the third and fourth switch valves 67 and 68 used for Exhaust and Intake VTCs 50 and 49 are identical in construction. Upon detecting the engine operation condition, crank angle and rotation angles of intake and exhaust drive shafts, the controller 22 derives or calculates a relative rotation angle between each of the exhaust and intake timing sprockets and each of the exhaust and intake camshafts 5 and 6.

In each of the intake and exhaust VTCs 49 and 50, between the vane member 52 and the housing 54, there is provided a lock mechanism that is able to lock and unlock the mechanical connection between the housing 54 and the vane member 52.

As will be understood from FIGS. 6A and 7, the lock mechanism is arranged between the larger vane 52 b of the vane member 52 and the rear cover 56 and comprises an elongate bore 72 that is formed in the larger vane 52, a cylindrical lock pin 73 that is slidably received in the elongate bore 72, a lock hole 74 a that is formed in a cup-shaped member 74 tightly fitted to an opening of the rear cover 56 and engageable with a tapered leading end 73 a of the lock pin 73, and a spring member 76 that is held by a spring retainer 75 fixed to a bottom part of the elongate bore 72 and biases the lock pin 73 toward the lock hole 74 a.

Although not shown in Hg. 7, there is provided an oil hole through which the hydraulic pressure is directly led into the lock hole 74 a from the retarding hydraulic pressure chambers 62 and/or from the oil pump 19.

When the vane member 52 is turned to the most retarded position (viz., default position), the tapered leading end 73 a of the lock pin 73 is brought into engagement with the lock hole 74 a due to the force of the spring member 76 locking the engagement between the timing sprocket 51 and the exhaust camshaft 6. However, when the hydraulic pressure is led into the lock hole 74 a from the retarding hydraulic pressure chambers 62 and/or from the oil pump 19, the lock pin 73 is moved back against the force of the spring member 76 canceling the locked engagement between the timing sprocket 51 and the exhaust camshaft 6. It is thus to be noted that the lock mechanism has a function to hold the default position (or condition). It is further to be noted that such lock mechanism may be removed.

As is seen from FIG. 6A, between the two circular recesses 52 c of each vane 52 b of the vane member 52 and adjacent one of the shoes 54 a of the housing 54, there are arranged two coil springs 77 and 78 for biasing the pared vane 52 b and shoe 54 a in opposed directions.

The two coil springs 77 and 78 are so arranged as not to contact each other even if the springs 77 and 78 are fully compressed. Although not shown in the drawing, a thin retainer is set between the bottom of the circular recess 52 c and one end of the coil spring 77 or 78.

In the following, operation of the exhaust VTC 50 will be described with reference to FIGS. 6A, 6B and 7. As will be mentioned hereinafter, operation of the intake VTC 49 is substantially the same as that of the exhaust VTC 50.

When the engine is stopped, the controller 22 stops feeding of a control signal to the third electromagnetic switch valve 67 causing the switch valve 67 to take the first position “I” (viz., retard control position) as shown in FIG. 7. In this condition, the passage 65 from the oil pump 19 and the second hydraulic pressure passage 64 are connected and at the same time the drain passage 66 and the first hydraulic pressure passage 63 are connected. Under this engine stop condition, the oil pump 19 is not energized and thus the hydraulic pressure from the oil pump 19 is 0 (zero).

Accordingly, as is seen from FIG. 6A, the vane member 52 is forced to take the most-retarded angular position due to the force of the coil springs 77 and 78. In this position, a left side (as viewed in FIG. 6A) of the larger vane 52 b is in contact with a right side of the corresponding shoe 54 a, and at the same time, the tapered leading end 73 a of the lock pin 73 of the lock mechanism is engaged with the lock h 74 a causing the vane member 52 to be locked to the most-retarded angular position. In other words, under such condition, the exhaust VTC 50 takes the mechanically stable default condition at the most-retarded angular position.

When, for starting the engine, an ignition switch is turned to ON position, the crankshaft of the engine is driven by a starter motor, the controller 22 starts to feed a control signal to the third electromagnetic switch valve 67. However, in an initial stage just after the cranking, the hydraulic pressure from the oil pump 19 fails to have a satisfied pressure and thus, the locked engagement between the vane member 52 and the housing 54 is still kept for a while keeping the stable default condition of the exhaust VTC 50 for a while.

Under this condition, the connection between the passage 65 and the second hydraulic pressure passage 64 and the connection between the drain passage 66 and the first hydraulic pressure passage 63 are kept.

When, due to the cranking, the hydraulic pressure from the oil pump 19 is somewhat raised, the raised hydraulic pressure is fed to the retarding hydraulic pressure chambers 62 through the second pressure passage 64. However, during this stage, the advancing hydraulic pressure chambers 61 are kept exposed to the oil pan 01 through the first hydraulic pressure passage 63 and the drain passage 66 receiving no hydraulic pressure from the oil pump 19.

When, in a later stage of the cranking, the hydraulic pressure from the oil pump 19 is sufficiently raised, the vane member 52 becomes freely controlled by the third electromagnetic switch valve 67 for the following reasons.

That is, in accordance with increase of the oil pump pressure and the pressure in the retarding hydraulic pressure chambers 62, the pressure in the lock hole 74 a of the lock mechanism is also increased to a certain level. Upon this, the lock pin 73 is moved back against the force of the spring member 76 and finally moved to a position to cancel the locked engagement between the vane member 52 and the housing 54. Upon this, the vane member 52 is permitted to turn (but small) freely relative to the housing 64.

For example, when, upon issuance of a certain signal from the controller 22, the third electromagnetic switch valve 67 takes the third position “III”, the passage 65 from the oil pump 19 is connected to the first hydraulic pressure passage 63 and at the same time, the drain passage 66 is connected to the second hydraulic pressure passage 64.

Under this condition, the hydraulic pressure in the retarding hydraulic pressure chambers 62 is drained to the oil pan 01 through the second hydraulic pressure passage 64 and the drain passage 66 causing the interior of the chambers 62 to show a lower pressure, and at the same time, the oil pump pressure is led into the advancing hydraulic pressure chambers 61 through the passage 65 and the first hydraulic pressure passage 63 causing the chambers 61 to show a higher pressure.

Upon this, as will be understood when comparing FIG. 6A and FIG. 6B, the vane member 52 is turned in a clockwise direction in the drawing against the force of the springs 77 and 78 toward the position as shown in FIG. 6B. Due to the clockwise turning of the vane member 52, the angular position of the exhaust camshaft 6 relative to the timing sprocket 51 is changed to an advanced side. When, due to control by the controller 22, the third electromagnetic switch valve 67 is sifted to the second position “II”, the vane member 52 can be stably held in any desired position relative to the housing 54.

In accordance with the engine operation condition after the warming up of the engine, the vane member 52 is continuously turned from the most-retarded position (FIG. 6A) to the most-advanced position (FIG. 6B).

Effects of the Second Embodiment

FIG. 8 shows a valve lift characteristic diagram depicting the valve lift characteristic of the intake valves 1 and 1 and exhaust valves 2 and 2, which is induced by the Intake VVL3, Exhaust VVL4, Intake VTC 49 and Exhaust VTC50 employed in the second embodiment of the present invention.

Like in the above-mentioned first embodiment, the default position provided by the Exhaust VVL4 causes a larger lift LE2 (larger operation angle DE2) and the default position provided by the Intake VVL3 causes a smaller lift LI1 (smaller operation angle DI1′). It is to be noted that the default operation angle DI1′ by the Intake VVL3 is set somewhat narrower than that DI1 of the first embodiment.

The default valve timings provided by the Intake and Exhaust VVLs 3 and 4 and the Intake and Exhaust VTCs 49 and 50 are depicted by the solid lift curves in FIG. 8. It is to be noted that the curves of the Intake and Exhaust VTCs are of the most-retarded default timing.

That is, in case of the Exhaust valves 2 and 2, the larger lift LE2 (larger operation angle DE2) is provided by the Exhaust VVL 4 and the most-retarded exhaust lift curve (solid line) is provided by the Exhaust VTC 50, and in case of the Intake valves 1 and 1, the smaller lift LI1 (smaller operation angle DI1′) is provided by the Intake VVL 3 and the most-retarded intake lift curve (solid line) is provided by the Intake VTC 49.

It is to be noted that the effect given by the above-mentioned feature as mentioned in the section of FIG. 8 is higher than that of the first embodiment. The reason is as follows.

In the second embodiment, the so-called default exhaust valve close timing (ECV) is retarded to the time EVC2′, and thus, the above-mentioned first period is increased to “(1)-2′” which is larger than “(1)-2” of the first embodiment, and thus, in the second embodiment, much larger amount of high temperature EGR gas can be led to the cylinder. Since, in the second embodiment, the valve open timing (IVO) of the default intake valves 1 and 1 is retarded to “IVO 1′”, the above-mentioned second period becomes smaller than that “(2)-1” (see FIG. 4) of the first embodiment. Thus, in the second embodiment, the timing of “IVO 1′” is placed very near the TDC. If necessary, as is seen from FIG. 8, the timing “IVO 1′” may be more retarded than TDC by the amount of “α”. That is, negative period may be provided.

If the valve timing is set in the above-mentioned manner, the intake valves 1 and 1 don't open even when the piston arrives at the top dead center in the exhaust stroke, and thus, exhaust gas that would be led back toward the intake side through the intake valves and thus cooled there is almost zero. In other words, re-intake of the lower temperature EGR gas (which is deteriorates combustion stability) into the cylinder is zero.

In the period of “α” in the intake stroke, the intake valves 1 and 1 are kept dosed and only the exhaust valves 2 and 2 are opened, and thus, in such period, downward moving of the piston does not intake fresh air except the high temperature EGR gas. In other words, the effect of introducing the high temperature EGR gas in the first period is much improved.

As is described hereinabove, in the second embodiment, much larger amount of high temperature EGR is led into the cylinder and almost zero of low temperature EGR is led into the cylinder in the limited period. Accordingly, combustion stability and exhaust emission at the time of engine starting are improved as compared with those of the first embodiment.

Because of the retarding control by the exhaust VTC 50, the open timing (EVO 2′) of the exhaust valves 2 and 2 is further retarded, and thus, the time for which the combustion gas is heated in the cylinder is increased, and thus, the time needed for warming up the engine in a cold start can be shortened.

Referring to FIG. 9, there is shown a flowchart that depicts controlled operation steps executed by the controller 22 of the second embodiment.

At step S21, judgment is carried out as to whether a current state is an engine stopping condition or not, that is, whether an engine key has returned to a key-off position or not.

If YES, the operation flow goes to step S22 where an engine stop signal (or fuel cut signal) is issued. Upon this, the engine speed Ne is gradually lowered. Due to lowering of the engine speed, the hydraulic pressure from the oil pump 19 is gradually lowered, and finally, the hydraulic pressure fails to have a mode changing force irrespective of the control positions of the electromagnetic switch valves 20 and 47 of the Intake and Exhaust VVL 3 and 4. Accordingly, due to the reasons as mentioned hereinabove, the intake and exhaust valves 1, 1, 2 and 2 are moved to their default positions as is mentioned at step S23.

That is, at step S23, the Intake and Exhaust VVLs are shifted to their default but mechanically stable conditions. The intake valves 1 and 1 take the default positions (viz., stable smaller operation angle) and the exhaust valves 2 and 2 take the default positions (viz., stable larger operation angle).

At the same time, in the Intake and Exhaust VTCs 49 and 50, the hydraulic pressure fails to have a mode changing force irrespective of the control positions of the fourth and third electromagnetic switch valves 68 and 67. Accordingly, due to the force of the coil springs 77 and 78 and a driving reaction torque of the valve actuating system, both the Intake and Exhaust VTCs 49 and 50 are shifted to their default but mechanically stable conditions (viz., the most-retarded conditions).

At step S24, operation of the engine is stopped.

At step S25, judgment is carried out as to whether the current state is an engine start condition or not, that is, whether the engine key is turned to a key-on position or not.

If YES, the operation flow goes to step S26 where, for assuring the positioning of the intake and exhaust valves 1, 1, 2 and 2 at the default positions, control or instruction signals are fed to the electromagnetic switch valves 20 and 47 of the Intake and Exhaust VVLs 3 and 4 as well as the electromagnetic switch valves 68 and 67 of the Intake and Exhaust VTCs 49 and 50 for enforcing the mechanisms 3, 4, 49 and 50 to take the above-mentioned default conditions. This control is aimed to deal with an undesirable event such as sudden change of the pump pressure during cranking, rotational fluctuation by the cranking or the like. Even in such case, the default conditions have to be kept by a small pump pressure.

At step S27, an instruction signal representing permission of starting the engine cranking is issued.

At step S28, judgment is carried out as to whether the cranking has arrived at a predetermined rotation speed or not.

If YES, the operation flow goes to step S29 where a fuel injection and ignition (viz., perfect ignition control) are made for starting the combustion (viz., starting combustion).

From an initial stage of this starting combustion, the above-mentioned default valve timings are kept, and thus, the starting combustion can be made stable thereby reducing the amount of the exhaust emission at the starting combustion. Furthermore, the mechanical fail safe function as mentioned in the section of the first embodiment is also obtained.

At step S30, judgment is carried out as to whether a predetermined time has passed from the engine starting or not. If YES, the operation flow goes to step S31 assuming that the initial engine rotation has come to a stable stage.

At step S31, an engine temperature T is derived from an information signal issued from a temperature sensor (not shown) mounted on a cylinder block.

At step S32, judgment is carried out as to whether the engine temperature T is lower than a predetermined temperature TO or not. If NO, the operation flow goes to step S33 assuming that the engine warming up operation is not needed any longer.

At step S33, instruction signals are fed to the Intake and Exhaust VVLs 3 and 4 and Intake and Exhaust VTCs 49 and 50 to control these mechanisms 3, 4, 49 and 50 in accordance with the engine speed Ne and data provided by a load map.

If YES at step S32, the operation flow goes to step S34 assuming that the engine warming up operation is further needed.

At step S34, target phases of the intake or exhaust valves 1 and 1 or 2 and 2 are calculated for each engine temperature.

If the engine temperature is still very low, such valves 1, 2, 2 and 2 are controlled in the above-mentioned default valve timing. That is, both the Intake and Exhaust VTCs 49 and 50 are controlled to take the most-retarded angular positions.

While, when the engine temperature shows a somewhat higher value, feeding a larger amount of the high temperature EGR gas to the cylinders of the engine tends to bring about abnormal combustion in the cylinders causing a knocking, pre-ignition or the like.

Accordingly, at step S35, a control for advancing the operation phase of the Exhaust VTC 50 is carried out in accordance with the engine temperature T from the most-retarded position (viz., default position).

With this control applied to the Exhaust VTC 50, the undesirable high temperature abnormal combustion can be suppressed by only advancing the operation phase of the Exhaust VTC 50 by a minimum degree that assures suppression of the abnormal combustion. With this suppression, combustion stability and lower emission in the period of engine starting are improved. This will be much easily understood from the valve lift characteristic diagram of FIG. 8. That is, first, the operation phase of the Exhaust VTC 50 is advanced to the exhaust valve lift characteristic of the first embodiment which is indicated by the broken line, and when thereafter the engine temperature is increased, the operation phase of the Exhaust VTC 50 is retarded to the advanced exhaust valve lift characteristic which is indicated by the dash-dot line.

If, in place of the Exhaust VTC 50, the operation phase of the Intake VTC 49 is controlled to be advanced from the most-retarded phase in accordance with the engine temperature, substantially the same effect is obtained.

When, as is mentioned hereinabove, the engine temperature exceeds the predetermined temperature T, the operation flow goes to step S33 estimating that the engine warming up operation is not needed any longer. At this step S33, a normal control of the Intake and Exhaust VVLs 3 and 4 and the Intake and Exhaust VTC 49 and 50 is started which is based on the engine speed Ne and data provided by the load map.

Third Embodiment

In FIG. 10, there is shown a variable valve system of a third embodiment of the present invention.

As will become apparent as the description proceeds, in this third embodiment, there is employed the Exhaust VVL 4 (used in the first embodiment) that uses a hydraulic pressure for changing the valve operation mode, and the maximum lift LE2 (maximum operation angle DE2) takes the default position.

However, the intake side is different from that of the first embodiment. That is, in the third embodiment, the intake side comprises an electric type Intake VEL 80 as an intake side lift varying mechanism and an electric type Intake VTC 81 as an intake side phase varying mechanism.

As will be understood from the drawing, the Intake VEL 80 is a mechanism that continuously varies the operation angle and lift amount of the intake valves 1 and 1, which is shown in for example Japanese Laid-open Patent Application (tokkai) 2012-225287.

For ease of understanding of the third embodiment, the Intake VEL 80 will be briefly described in the following with reference to accompanying drawings FIGS. 10 to 12. For simplification of the description, detailed description on the Exhaust VVL 4 will be omitted leaving in the drawings the same reference numerals of parts as those of the first embodiment.

As will be seen from FIGS. 11A and 11B, due to operation of the Intake VEL 80, the intake valves 1 and 1 can assume the minimum lift LI0 (minimum operation angle DI0) and the maximum lift LI2 (maximum operation angle DI2).

As is shown in FIG. 10, the Intake VEL 80 comprises a hollow driving shaft 82 that is rotatably supported by bearings provided on a cylinder head (not shown), a rotational cam 83 that is tightly mounted on the driving shaft 82, two swing cams 84 and 84 that are rotatably supported by the driving shaft 82 and contactable to upper surfaces of valve lifters 1 a and 1 a to cause the intake valves 1 and 1 to make their opening action, and a transmission mechanism that is arranged between the rotational cam 83 and each of the swing cams 84 and 84 to convert the rotation of the rotational cam 83 to swing movements of the swing arms 84 and 84.

The drive shaft 82 has at one end a timing sprocket 85 around which a timing belt (not shown) from the crankshaft is put for transmitting rotation of the crankshaft to the drive shaft 82. In FIG. 10, a rotation direction of the drive shaft 82 is indicated by an arrow.

As will be understood from FIGS. 10 and 11A, the rotational cam 83 has a generally ring-like shape. However, as is seen from FIG. 11, upon assembly, a shaft center Y of the rotational cam 83 is offset in a radial direction relative to a shaft center X of the drive shaft 82. Although not well shown in FIG. 10, the two swing arms 84 are fixed to opposed ends of a cam shaft (not shown) that is rotatably disposed on the drive shaft 82. Thus, the two swing arms 84 swing like one unit. Each swing cam 84 has a cam surface including a base circle surface, a ramp surface and a lift surface. In operation, the base circle, ramp surface and lift surface of the swing cam 84 are brought into contact in order with a given portion of the corresponding valve lifter 1 a in accordance with swing movement of the swing cam 84.

As is seen from FIG. 10, the transmission mechanism comprises a rocker arm 86, a link arm 87 and a link rod 88. One end of the rocker arm 86 is pivotally connected to the link arm 87, and the other end of the rocker arm 86 is pivotally connected to one end of the link rod 88, as shown.

As shown, the link arm 87 has a circular bore in which a cam proper of the rotational cam 83 is rotatably received. The link arm 87 has further a projected part rotatably connected to one end of the rocker arm 86 through a pin (no numeral). The other end of the link rod 88 is pivotally connected to a cam nose of one of the swing cams 84 through a pin.

As will be understood from FIG. 10, above the drive shaft 82, there is arranged a control shaft 89 that is rotatable about its axis. A control cam 90 is tightly mounted on the control shaft 89 to serve as a rocking fulcrum of the rocker arm 86. A rotation of the control shaft 89 is controlled by an electric actuator 91. As shown, a shaft center P2 of the control cam 90 is offset by a given degree relative to a shaft center P1 of the control shaft 89.

As is seen from FIGS. 12A and 12B, the electric actuator 91 comprises a casing 91 a, an electric motor 92 fixed to one end of the casing 91 a, and a ball-screw transmission mechanism 93 that is installed in the casing 91 a to transmit a rotational driving force of the electric motor 92 to the control shaft 89.

The electric motor 92 is of a proportional DC type and controlled by the controller 22 in accordance with an operation condition of the engine.

As is seen from FIGS. 10 and 12A, the ball-screw transmission mechanism 93 comprises a ball-screw shaft 93 a that is arranged coaxial with a drive shaft of the electric motor 92, a ball nut 93 b that is disposed on the ball-screw shaft 93 a while making a meshed engagement therebetween, and a connecting arm 93 c that is connected to one end of the control shaft 89 and has one projected end pivotally connected to the ball nut 93 b through a link (see FIG. 10).

As is seen from FIG. 10, the ball-screw shaft 93 a has one end fixed to the drive shaft of the electric motor 92, and thus, the ball-screw shaft 93 a is rotated about its axis by the motor 92.

As shown in FIG. 12A, the ball nut 93 b is shaped cylindrical and has on its inner cylindrical surface a spirally extending guide groove for rotatably receiving a plurality of balls in cooperation with ball circulation grooves. With this arrangement, the ball nut 93 b is moved axially on and along the ball-screw shaft 93 a while changing the rotational movement of the ball-screw shaft 93 a to the axial movement thereof. As is seen from FIGS. 10 and 12A, the ball nut 93 b is biased by a coil spring 94 toward the electric motor 92, that is, in a direction to provide the intake valves 1 and 1 with the minimum lift. Accordingly, when the engine is stopped, the coil spring 94 moves the ball nut 93 b to the position that provides the minimum lift of the intake valves 1 and 1. That is, the default position of the Intake VEL 80 brings about the minimum lift LI0 (minimum operation angle DI0).

The controller 22 used in this third embodiment is substantially the same as the controllers 22 used in the first and second embodiments except the following.

That is, in this third embodiment, by processing an information signal from a drive shaft angle sensor (not shown) detecting a rotation angle of the drive shaft 82 and an information signal from a potentiometer 95 detecting a rotational position (displacement) of the control shaft 89, the controller 22 outputs instruction signals for controlling both the electric type Intake VTC 81 to cause the drive shaft 82 to take a certain rotational angle relative to the crank angle and the electric type Intake VEL 80 to cause the intake valves 1 and 1 to take a certain valve lift amount and a certain operation angle.

Accordingly, when, in a given operation zone, the electric motor 92 is driven or turned in a given direction upon receiving instruction signal (or current) from the controller 22, the ball-screw shaft 93 a is turned in a given direction. Upon this, as is seen from FIG. 12A, the ball nut 93 a is linearly moved toward the electric motor 92 with the help of the coil spring 94, which turns the control shaft 89 in a given direction through the connecting arm 93 c.

Accordingly, as will be understood from FIG. 11A, the control cam 90 fixed to the control shaft 89 is turned about the shaft center of control shaft 89 while shifting a thicker part thereof upward away from the drive shaft 82. With this, the other end of the rocker arm 86 and the pivot point of the link rod 88 are moved upward relative to the drive shaft 82 and thus the cam noses of the swing cams 84 are moved upward through the link rod 88 causing the swing arms 84 to turn in a clockwise direction in FIG. 11A.

When, due to rotation of the rotational cam 83, the end portion of the rocker arm 86 is lifted through the link arm 87, the lift amount (viz., shift) of the end portion is transmitted to the swing cams 84 and valve lifters 1 a (see FIG. 11A) through the link rod 88, so that as is seen from the graph of FIG. 13, the intake valves 1 and 1 take a movement of the minimum lift (LI0) and minimum operation angle (DI0).

When the engine is shifted to a high speed high load operation range, the controller 22 controls the electric motor 92 to turn in a reversed direction causing the ball nut 93 b to move to the rightmost position as is shown in FIG. 12B. With this, as is seen from FIGS. 11A and 11B, the control shaft 89 rotates the control cam 90 in a clockwise direction moving the shaft center P2 of the control cam 90 downward. Accordingly, as is seen from FIG. 11B, the rocker arm 86 is entirely shifted toward the drive shaft 82. During this, the other end of the rocker arm 86 presses down the nose part of one of the united swing cams 84 through the link rod 88 turning the swing cams 84 in a counterclockwise direction by a certain angle.

Accordingly, when thereafter, due to rotation of the rotational cam 83, the end portion of the rocker arm 86 is lifted through the link arm 87, the lift amount (shift) of the end portion is transmitted to the swing cams 84 and the valve lifters 1 a in the above-mentioned manner. However, in this case, as is seen from the graph of FIG. 13, the lift of the intake valves 1 and 1 is gradually increased like the sequence of “LI1-LI1.5-LI2”. As a result, the exhaust efficiency of the engine in the high speed area is increased and thus increased output is provided.

That is, as is understood from the above, the lift amount of the intake valves 1 and 1 is continuously varied from the minimum lift LI0 to the maximum lift LI2, and thus, also the operation angle of the intake valves 1 and 1 is continuously varied from the minimum operation angle DI0 to the maximum operation angle DI2.

When the engine is stopped, the ball nut 93 b is moved toward the electric motor 92 to take the certain position due to the force of the coil spring 94 as is mentioned hereinabove. Accordingly, the minimum operation angle DI0 and the minimum lift LI0 (viz., default position) of the intake valves 1 and 1 are stably held by the mechanism.

Since the electric type Intake VTC 81 is described in Japanese Laid-open Patent Application (tokkai) 2012-145036, the detailed explanation of its construction will be omitted. In the Intake VTC 81, the rotational phase of the drive shaft 82 is varied by a force of the electric motor through a speed reduction mechanism. Furthermore, a biasing spring (not shown) is used which biases the drive shaft 82 to rotate in a retarding direction. Thus, the default position of the Intake VTC 81 takes the most retarded angle like in case of the second embodiment.

Accordingly, when controlled to take their default position by the Intake VEL 80 and Intake VTC 81, the intake valves 1 and 1 show the lift curve of LI0 in FIG. 13.

Thus, the combustion stability and emission reduction at the time of engine starting are much improved as compared with those of the above-mentioned first and second embodiments.

The reason of the superiority of the third embodiment is as follows.

When the intake valves 1 and 1 take the default position, the opening timing (IVO) of the intake valves 1 and 1 shows the time point “IVO0” which is more retarded than that of the second embodiment. Actually, in the third embodiment, the opening timing (IVO) is retarded to that of the second embodiment by the amount “α”, and thus, the above-mentioned “second period” disappears, and thus even when the piston comes up to the top dead center in the exhaust stroke, the intake valves 1 and 1 don't open. Accordingly, the phenomenon wherein the combustion gas is led back toward the intake side through the intake valves 1 and 1 to be cooled and then led into the cylinder again as a low temperature EGR gas is not provided.

Furthermore, since, in the period of the amount “α”, the intake valves 1 and 1 are closed and only the exhaust valves 2 and 2 are opened, and thus, when, in the period, the piston is moved down, only the high temperature EGR gas is led into the cylinder without sucking a fresh air. Accordingly, the high temperature EGR gas introducing effect is much raised, which exceeds the effect of the second embodiment.

In the third embodiment, the lift amount (or operation angle) of the intake valves 1 and 1 in the default condition is sufficiently reduced and the closing timing (IVCO) of the intake valves 1 and 1 is set near the bottom dead center, the amount “α” has a sufficient amount and thus the above-mentioned effect is more improved.

Furthermore, in this third embodiment, the lift amount LI0 is sufficiently small, and thus air intake speed is increased, which promotes the air/fuel mixture mixing effect in the cylinder.

Furthermore, in the third embodiment, between the exhaust valve closing timing (EVC2) and the intake valve opening timing (IVO0), there is provided a period “β” for which all the intake and exhaust valves 1, 1, 2 and 2 are closed. In the period “β”, there is a growth of negative pressure in the cylinder by a downward movement of the piston, and thus, the air intake speed at the time of opening the intake valves 1 and 1 is increased and thus combustion of the air/fuel mixture in the cylinder (viz., combustion chamber) is much improved.

In the third embodiment, the closing timing of the intake valves 1 and 1 is set near the bottom dead center. Thus, the third embodiment has merits provided by the improvement of the effective compression ratio like in the first and second embodiments.

Furthermore, in the third embodiment, the exhaust valves 2 and 2 take the larger operation angle DE2 (larger lift LE2), and thus, the third embodiment has an improved combustion like in the first and second embodiments.

However, since, in the third embodiment, the combustion improvement effect is high, increasing the temperature of the engine tends to induce undesired knocking or pre-ignition of the engine. Thus, for eliminating such undesired phenomena, the following control is carried out. That is, by the control, the operation angle of the intake valves 1 and 1 by the Intake VEL 80 is finely controlled, and at the same time the operation of the intake valves 1 and 1 by the Intake VTC 81 is finely advanced.

Referring to FIG. 14, there is shown a flowchart that depicts controlled operation steps executed by the controller 22 of the third embodiment. That is, as will become apparent as the description proceeds, the flowchart shows the operation steps for controlling the Intake VEL 80, Intake VTC 81 and Exhaust VVL 4.

At step S41, judgment is carried out as to whether a current state is an engine stopping condition or not, that is, whether an engine key has returned to a key-off position or not.

If YES, the operation flow goes to step S42 where an engine stop signal (or fuel cut signal) is issued. Upon this, the engine speed Ne is gradually lowered. Due to lowering of the engine speed, the hydraulic pressure from the oil pump 19 is gradually lowered, and finally, the hydraulic pressure fails to have a mode changing force irrespective of the control position of the electromagnetic switch valve 47 of the hydraulically actuated Exhaust VVL 4.

At step S43, the exhaust valves 2 and 2 are forced to take the default condition (viz., larger lift LE2, larger operation angle DE2) which is stable. The electric Intake VEL 80 and the electric Intake VTC 81 are de-energized, and thus, the intake valves 1 and 1 are gradually shifted to the default condition when the minimum lift LI0 of the valves 1 and 1 and the most-retarded operation phase are established.

At step S44, operation of the engine is stopped.

At step S45, judgment is carried out as to whether the current state is an engine start condition or not, whether the engine key has be turned to a key-on position or not.

If YES, the operation flow goes to step S46 where control signals are fed to the Exhaust VVL 4, Intake VEL 80 and Intake VTC 81 to force these mechanisms 4, 80 and 81 to take the default condition. This step is aimed to deal with an undesirable event such as sudden change of the pump pressure during cranking, rotational fluctuation by the cranking or the like. Even in such case, the default condition has to be kept by a small pump pressure.

At step S47, an instruction signal representing permission of starting the engine cranking is issued.

At step S48, judgment is carried out as to whether the cranking has arrived at a predetermined rotation speed or not.

If YES, the operation flow goes to step S49 where a fuel injection and ignition (viz., perfect ignition control) are made for starting the combustion (viz., starting combustion).

From an initial stage of this starting combustion, the above-mentioned default valve timings are kept, and thus, the starting combustion can be made stable thereby reducing the amount of the exhaust emission at the starting combustion. Furthermore, the mechanical fail safe function as mentioned in the first and second embodiments is also obtained.

At step S50, judgment is carried out as to whether a predetermined time has passed from the engine starting or not. If YES, the operation flow goes to step S51 judging that the initial engine rotation has come to a stable stage.

At step S51, an engine temperature T is derived from an information signal issued from a temperature sensor (not shown) mounted on a cylinder block.

At step S52, judgment is carried out as to whether the engine temperature T is lower than a predetermined temperature TO or not. If NO, the operation flow goes to step S53 assuming that the engine warming up operation is not needed any longer.

At step S53, instruction signals are fed to the Intake VEL 80, Exhaust VVL 4 and Intake VTC 81 to control these mechanisms 80, 4 and 81 in accordance with the engine speed Ne and data provided by a load map.

If YES at step S52, the operation flow goes to step S54 assuming that the engine warming up operation is further needed.

At step S54, target instruction values that are to be provided by the Intake VEL 80 and Intake VTC 81 are calculated for each engine temperature.

At step S55, the Intake VEL 80 and Intake VTC 81 are finely controlled based on the calculated target instruction values.

If the engine temperature is still very low, the intake valves 1 and 1 are controlled in the above-mentioned default valve timing.

While, when the engine temperature shows a somewhat higher value, feeding a larger amount of the high temperature EGR gas to the cylinders of the engine tends to bring about abnormal combustion in the cylinder causing knocking, pre-ignition or the like.

Accordingly, at step S55, a control for gradually advancing the operation phase of the Intake VEL 80 from the minimum lift LI0 (minimum operation angle DI0) toward LI2 (see FIG. 13) through lift LI1 and lift LI1.5 is carried out in accordance with the engine temperature T, and at the same time, the operation phase provided by the Intake VTC 81 is gradually advanced keeping the close timing of the intake valves 1 and 1 near the bottom dead center. With this action, combustion stability and emission reduction can be obtained while suppressing the high temperature abnormal combustion in the cylinders.

Accordingly, when the engine temperature exceeds the predetermined temperature TO, the Intake VEL 80, Exhaust VVL 4 and Intake VTC 81 are controlled to take their normal operation modes assuming that the warming up operation of the engine is not needed any longer. The normal operation modes are carried out based on the engine speed Ne and the load map.

For example, when the operation of the engine is shifted to the high speed-high load range, the Intake VEL 4 changes the lift of the intake valves 1 and 1 to the maximum lift LI2 and the Intake VTC 81 shifts the operation phase toward a retarded side as is seen from the broken line in FIG. 13, so that the close timing of the intake valves 1 and 1 is retarded to the timing IVC2 while controlling the change (≈IVO1.5) of the open timing IVO2 of the valves 1 and 1. With this, the air/fuel mixture charging efficiency in the high speed range is increased while suppressing a change of the amount of EGR gas, so that the output of the engine is stably and smoothly increased.

When, in this third embodiment, the engine temperature is still below TO due to insufficient warming up operation of the engine, the Intake VEL 80 and Intake VTC 81 which are powered by electricity are preferentially actuated, and thus, the intake valves 1 and 1 can have a high responsibility.

Furthermore, as is seen from FIG. 10, in the third embodiment, the power source for actuating the variable valve system includes an electric power that actuates the Intake VEL 80 and Intake VTC 81 and a hydraulic power that actuates the Exhaust VVL 4. Now, let us assume that the above-mentioned Intake and Exhaust systems are powered by only the hydraulic power. In this case, an oil pump is usually used which means necessity of providing two or more hydraulic passages for the two systems. Thus, the possibility of undesirable oil leakage is inevitably increased, and when the two systems are actuated at the same time, the pressure feeding speed of the oil pump is reduced, which deteriorates the control responsiveness.

However, in the third embodiment, only the Exhaust VVL 4 is powered by the hydraulic power, and thus, the above-mentioned drawbacks can be avoided or at least minimized.

Now, let us assume that the Intake and Exhaust valve control systems are powered by only the electric power. In this case, the battery load is increased and furthermore, a problem arises in the layout of the variable valve system. That is, like the electric actuator of the Intake VEL 80 which has larger mounting spaces for the electric motor 92 and the speed reduction mechanism, the electric actuator of a mechanism that serves as the Exhaust VVL 4 has to prepare larger mounting spaces for an electric motor and a speed reduction mechanism. Accordingly, mountability of such systems, which are bulky in construction, to the engine room is deteriorated. While, when, like in the third embodiment, the Intake and Exhaust valve control systems are respectively powered by the electric power and the hydraulic power, the mountability is improved.

It is to be noted that the present invention is not limited to the above-mentioned three embodiments. Some modifications and variations of such embodiments will be briefly described in the following.

For example, the lift varying mechanism and the phase varying mechanism may be of a type that can carry out a two step change or a continuous change. In such case, the power for actuating the mechanisms may be a hydraulic power, electric power or pneumatic power.

In the above-mentioned embodiments, a coil (or biasing) spring is used for causing the lift/phase varying mechanisms to take the default conditions when no power is applied thereto from the power source (or power sources). However, if necessary, in place of the coil spring, a power provided by the valve actuating system may be used. Furthermore, a lock pin system may be employed for obtaining the stable default position or condition of the lift/phase varying mechanisms.

In the above-mentioned embodiments, the valve open/close timings (IVO, IVC, EVO and EVC) of the intake and exhaust valves 1, 1, 2 and 2 are determined based on just the times when the lift of the valves 1 and 2 starts and ends. However, if desired, such valve open/close timings may be determined with reference to a so-called ramp period that appears at an initial stage of the lift and an end stage of the lift. In the latter case, since the gas exchanging effectively starts and effectively ends, the effects of the present invention are much improved.

Furthermore, if the technology disclosed by Japan Laid-open Patent Application (tokkai) 2002-276446 is practically applied to the present invention, the effects of the present invention are much improved. That is, the publication shows a technology for adjusting the top dead center of a piston. If such technology is employed, a so-called geometric compression ratio is increased and thus the effective compression ratio can be increased, resulting in that the combustion effect at the engine starting is much more improved.

The entire contents of Japanese Patent Application 2014-56299 filed Mar. 3, 2014 are incorporated herein by reference.

Although the invention has been described above with reference to the embodiments of the invention, the invention is not limited to the such embodiments as described above. Various modifications and variations of such embodiments may be carried out by those skilled in the art in light of the above description. 

What is claimed is:
 1. A variable valve system of an internal combustion engine, comprising: an intake valve operation angle varying mechanism that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto; and an exhaust valve operation angle varying mechanism that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, wherein the intake valve operation angle varying mechanism forces mechanically the intake valve to take a smaller operation angle when receiving no conversion actuation force; and wherein the exhaust valve operation angle varying mechanism forces mechanically the exhaust valve to take a larger operation angle when receiving no conversion actuation force.
 2. A variable valve system of an internal combustion engine as claimed in claim 1, further comprising: an exhaust valve operation phase varying mechanism that varies an exhaust valve operation phase at the time when the exhaust valve assumes its peak lift, wherein the exhaust valve operation phase varying mechanism forces mechanically the exhaust valve operation phase to take a retarded side when receiving no conversion actuation force.
 3. A variable valve system of an internal combustion engine as claimed in claim 1, further comprising: an intake valve operation phase varying mechanism that varies an intake valve operation phase at the time when the intake valve assumes its peak lift, wherein the intake valve operation phase varying mechanism forces mechanically the intake valve operation phase to take a retarded side when receiving no conversion actuation force.
 4. A variable valve system of an internal combustion engine as claimed in claim 1, further comprising: an intake valve operation phase varying mechanism that varies an intake valve operation phase at the time when the intake valve assumes its peak lift; and an exhaust valve operation phase varying mechanism that varies an exhaust valve operation phase at the time when the exhaust valve assumes its peak lift, wherein the intake and exhaust valve operation phase varying mechanisms force mechanically the intake and exhaust valve operation phases to take a retarded side when receiving no conversion actuation forces respectively.
 5. A variable valve system of an internal combustion engine as claimed in claim 1, in which at least one of the conversion actuation forces respectively applied to the intake and exhaust valve operation angle varying mechanisms is a hydraulic force.
 6. A variable valve system of an internal combustion engine as claimed in claim 1, in which at least one of the conversion actuation forces respectively applied to the intake and exhaust valve operation angle varying mechanisms is an electric force.
 7. A variable valve system of an internal combustion engine as claimed in claim 1, in which one of the conversion actuation forces respectively applied to the intake and exhaust valve operation angle varying mechanisms is a hydraulic force and the other of the conversion actuation forces is an electric force.
 8. A variable valve system of an internal combustion engine as claimed in claim 1, in which: when receiving the conversion actuation force, the intake valve operation angle varying mechanism forces the intake valve to take an operation angle larger than that taken when receiving no conversion actuation force.
 9. A variable valve system of an internal combustion engine as claimed in claim 8, in which: when receiving the conversion actuation force, the exhaust valve operation angle varying mechanism forces the exhaust valve to take an operation angle smaller than that taken when receiving no conversion actuation force.
 10. A variable valve system of an internal combustion engine as claimed in claim 9, in which: the intake and exhaust valve operation angle varying mechanisms selectively vary the respective operation angles of the intake and exhaust valves based on a fact whether or not the conversion actuation forces are applied to the mechanisms respectively.
 11. A variable valve system of an internal combustion engine as claimed in 9, in which: the intake and exhaust valve operation angle varying mechanisms continuously vary the respective operation angles of the intake and exhaust valves in accordance with an operation amount of the conversion actuating forces respectively.
 12. A variable valve system of an internal combustion engine, comprising: an intake valve operation angle varying mechanism that varies an operation angle of an intake valve by using a hydraulic force or electric force as a conversion actuation force; and an exhaust valve operation angle varying mechanism that varies an operation angle of an exhaust valve by using a hydraulic force or electric force as a conversion actuation force; wherein when receiving no conversion actuation force, the intake valve operation angle varying mechanism forces the intake valve to take and hold an operation angle smaller than that taken when receiving the conversion actuation force, and wherein when receiving no conversion actuation force, the exhaust valve operation angle varying mechanism forces the exhaust valve to take and hold an operation angle larger than that taken when receiving the conversion actuation force.
 13. A variable valve system of an internal combustion engine as claimed in claim 12, further comprising: an exhaust valve operation phase varying mechanism that varies an exhaust valve operation phase at the time when the exhaust valve assumes its peak lift, wherein the exhaust valve operation phase varying mechanism forces mechanically the exhaust valve operation phase to take a retarded side when receiving no conversion actuation force.
 14. A variable valve system of an internal combustion engine as claimed in claim 12, further comprising: an intake valve operation phase varying mechanism that varies an intake valve operation phase at the time when the intake valve assumes its peak lift, wherein the intake valve operation phase varying mechanism forces mechanically the intake valve operation phase to take a retarded side when receiving no conversion actuation force.
 15. A variable valve system of an internal combustion engine as claimed in claim 12, further comprising: an intake valve operation phase varying mechanism that varies an intake valve operation phase at the time when the intake valve assumes its peak lift; and an exhaust valve operation phase varying mechanism that varies an exhaust valve operation phase at the time when the exhaust valve assumes its peak lift, wherein the intake and exhaust valve operation phase varying mechanisms force mechanically the intake and exhaust valve operation phases to take a retarded side when receiving no conversion actuation force.
 16. A variable valve system of an internal combustion engine, comprising: a first means that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto; and a second means that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, wherein the first means forces mechanically the intake valve to take a smaller operation angle when receiving no first conversion actuation force; and wherein the second means forces mechanically the exhaust valve to take a larger operation angle when receiving no second conversion actuation force.
 17. In a variable valve system of an internal combustion engine which comprises an intake valve operation angle varying mechanism that varies an operation angle of an intake valve in accordance with a conversion actuation force applied thereto and an exhaust valve operation angle varying mechanism that varies an operation angle of an exhaust valve in accordance with the conversion actuation force applied thereto, a method for controlling the variable valve system comprising: controlling the intake valve operation angle varying mechanism to force mechanically the intake valve to take a smaller operation angle when no conversion actuation force is applied to the intake valve operation angle varying mechanism; and, controlling the exhaust valve operation angle varying mechanism to force mechanically the exhaust valve to take a larger operation angle when no conversion actuation force is applied to the exhaust valve operation angle varying mechanism. 